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United States Patent |
5,080,642
|
Takami
|
January 14, 1992
|
Rotation transmission device with a torque limiting transmission gear
mechanism
Abstract
A rotation transmission device with a torque limiting function comprises a
cylindrical housing 10, a first rotation member 20 supported on a first
bracket 12 via a one-way clutch bearing 14, a second rotation member 30
integral with the output shaft 33, an annular torsional resilient member
80, and an input shaft 40 rotatably supported on the rotation members 20
and 30. The torque is transmitted from the input shaft to the first
rotation member via a torque limiting transmission mechanism which
comprises: a first and a second internal gear 21 and 31, a first sun gear
41 supported on the input shaft 40 via a one-way clutch bearing 42, a
second sun gear 43 fixed on the input shaft, first and second planetary
gears 55 and 65, satellite rings 58 and 68 eccentric to the planetary
gears 55, and a guide disk 70 having guide slots 72 slidably engaging with
the satellite rings 58 and 68. The first rotation member 20 is driven and
rotated by the torque limiting transmission mechanism relative to the
second rotation member 30 while the torsional torque of the resilient
member 80 is under a predetermined maximum. When the torsional torque of
the resilient member 80 reaches the maximum the torque limiting
transmission mechanism rotates freely without transmitting torque to the
first rotation member 20.
Inventors:
|
Takami; Akira (Himeji, JP)
|
Assignee:
|
Mitsubishi Denki Kabushiki Kaisha (Tokyo, JP)
|
Appl. No.:
|
615959 |
Filed:
|
November 20, 1990 |
Foreign Application Priority Data
| Nov 20, 1989[JP] | 1-302597 |
| Nov 28, 1989[JP] | 1-310387 |
Current U.S. Class: |
475/263; 475/318; 475/330; 475/347 |
Intern'l Class: |
F16D 043/20 |
Field of Search: |
475/263,317,318,330,346,347
|
References Cited
U.S. Patent Documents
4173059 | Nov., 1979 | Hashimoto et al. | 29/240.
|
4869139 | Sep., 1989 | Gotman | 81/475.
|
Primary Examiner: Braun; Leslie A.
Assistant Examiner: Trousdell; William O.
Attorney, Agent or Firm: Sughrue, Mion, Zinn, Macpeak and Seas
Claims
What is claimed is:
1. A rotation transmission device for transmitting a torque in a
predetermined rotational direction with a torque limiting function,
comprising:
a hollow cylindrical housing having two ends closed by a first and a second
bracket, respectively;
a first rotation member concentrically disposed within said housing and
supported by said first bracket of the housing such that it is rotatable
only in said predetermined rotational direction;
an output shaft rotatably supported by said second bracket of the housing;
a second rotation member concentrically disposed within said housing and
connected integrally to said output shaft and rotatably supported by said
second bracket of the housing, said first and second rotation members
opposing each across an axial length within said housing;
an annular torsional resilient member connecting the first and the second
rotation members across said axial length, said torsional resilient member
exerting between the first and the second rotation members a torsional
torque proportional to a relative rotational displacement of the first
rotation member with respect to the second rotation member;
an input shaft driven in said predetermined rotational direction and
concentrically extending within said first and second rotation members to
be rotatably supported by said first and second rotation members; and
a torque transmission mechanism for transmitting torque from the input
shaft to the first rotation member while the torsional torque exerted by
said torsional resilient member from the first to the second rotation
member in said predetermined rotational direction is below a predetermined
magnitude, said torque transmission mechanism limiting under the
predetermined magnitude the torque transmitted from the first to the
second rotation member.
2. A rotation transmission device as claimed in claim 1, wherein said
torque transmission mechanism comprises:
a first internal gear formed integrally with said first rotation member to
extend axially therefrom toward the second rotation member;
a second internal gear formed integrally with said second rotation member
to extend axially therefrom toward the first rotation member;
a first sun gear supported concentrically on said input shaft in axial
alignment with said first internal gear;
a second sun gear supported concentrically on said input shaft in axial
alignment with said second internal gear, wherein: the first sun gear is
supported on the input shaft such that it is rotatable only in a direction
opposite to said predetermined rotational direction while the second sun
gear is fixed on the input shaft;
a plurality of first planetary gears meshing with said first sun gear and
first internal gear and rotatably supported at an equal eccentricity with
respect to the input shaft on a first carrier rotatably supported on the
input shaft, such that the first planetary gears are capable of planetary
motion around the first sun gear;
a plurality of second planetary gears meshing with said second sun gear and
second internal gear and rotatably supported at an equal eccentricity with
respect to the input shaft on a second carrier rotatably supported on the
input shaft, such that the second planetary gears are capable of planetary
motion around the second sun gear;
first satellite shafts secured on and extending axially from respective
first planetary gears with an eccentricity with respect to respective axes
of rotation of the first planetary gears;
second satellite shafts secured on and extending from respective second
planetary gears with an eccentricity with respect to respective axes of
rotation of the second planetary gears, the eccentricity of the second
satellite shafts with respect to the respective axes of the second
planetary gears being equal to the eccentricity of the first satellite
shafts with respect to the respective axes of the first satellite gears;
and
a disk-shaped guide member rotatably supported on the input shaft between
said first and second rotation members and having radially extending guide
slots slidably engaging with said first and second satellite shafts, such
that a torque in said predetermined direction is transmitted via the guide
member while the torque is below said predetermined magnitude.
3. A rotation transmission device as claimed in claim 2, wherein said guide
slots of the guide member engage with said first and second satellite
shafts via satellite rings rotatably and concentrically supported on the
satellite shafts.
4. A rotation transmission device as claimed in claim 2, wherein
revolutional angles of the first satellite shafts around respective axes
of the first planetary gears are initially different from revolutional
angles of the second satellite shafts around respective axes of the second
planetary gears.
5. A rotation transmission device as claimed in claim 1, wherein said
torque transmission mechanism comprises:
a first sun gear concentrically supported on said input shaft;
a second sun gear concentrically supported on said input shaft, the first
sun gear is supported on the input shaft such that it is rotatable only in
a direction opposite to said predetermined rotational direction while the
second sun gear is fixed on the input shaft;
a plurality of first planetary gears meshing with said first sun gear and
rotatably supported at an equal eccentricity with respect to the input
shaft on a first carrier formed integrally with said first rotation
member, such that the first planetary gears are capable of planetary
motion around the first sun gear;
a plurality of second planetary gears meshing with said second sun gear and
rotatably supported at an equal eccentricity with respect to the input
shaft on a second carrier formed integrally with said second rotation
member, such that the second planetary gears are capable of planetary
motion around the second sun gear;
first satellite shafts secured on and extending axially from respective
first planetary gears with an eccentricity with respect to respective axes
of rotation of the first planetary gears;
second satellite shafts secured on and extending from respective second
planetary gears with an eccentricity with respect to respective axes of
rotation of the second planetary gears, the eccentricity of the second
satellite shafts with respect to the respective axes of the second
planetary gears being equal to the eccentricity of the first satellite
shafts with respect to the respective axes of the first planetary gears;
and
a disk-shaped guide member rotatably supported on the input shaft between
said first and second rotation members and having radially extending guide
slots slidably engaging with said first and second satellite shafts, such
that a torque in said predetermined direction is transmitted via the guide
member while the torque is below said predetermined magnitude.
6. A rotation transmission device as claimed in claim 5, wherein said guide
slots of the guide member engage with said first and second satellite
shafts via satellite rings rotatably and concentrically supported on the
satellite shafts.
7. A rotation transmission device as claimed in claim 5, wherein
revolutional angles of the first satellite shafts around respective axes
of the first planetary gears are initially different from revolutional
angles of the second satellite shafts around respective axes of the second
planetary gears.
Description
BACKGROUND OF THE INVENTION
This invention relates to a rotation transmission device for transmitting a
torque with a torque limiting function, and more particularly to such a
device with a purely mechanical torque limiting transmission mechanism.
Conventionally, torque limiting transmission devices utilize, as the torque
limiting transmission means, friction, viscosity or dynamic pressure of a
fluid, or an electromagnetic force. While the magnitude of the load acting
on the output shaft is below a predetermined maximum, a torque equal to
the load is tranmitted from the input to the output shaft. However, when
the load exceeds the predetermined maximum, the input shaft slips relative
to the output, thereby limiting the torque of the output shaft under the
predetermined maximum.
The conventional torque limiting transmission devices have the following
characteristics: First, throughout during the torque limiting mode
operation, the input shaft slips relative to the input shaft. Second, also
throughout the torque limiting mode operation, the torque equal to the
torque limiting value (the predetermined maximum) act on the input as well
as the output shaft.
Thus, the conventional device has the following problems. Namely, within
the torque limiting transmission is generated a power loss equal to the
product of the predetermined maximum torque limiting value and the
relative slipping rotational velocity between the input and the output
shafts. Almost all the power loss is converted into thermal energy, and
raises the temperature of the component parts in the neighborhood. Thus,
measures should be taken to prevent the deformation or degeneration of the
parts of the torque limiting transmission means. Further, there arises the
problem of the function or performance that, due to the temperature rise,
the maximum torque limiting value or the operational durability vary
unstably.
SUMMARY OF THE INVENTION
It is therefore a primary object of this invention to provide an
inexpensive and purely mechanically organized rotation transmission device
with a torque limiting function wherein no power loss takes place when the
load is above the predetermined maximum limit, thereby enhancing the
reliability and durability.
The above object is accomplished according to the principle of this
invention by a rotation transmission device for transmitting a torque in a
predetermined rotational direction with a torque limiting function, which
device comprises: a hollow cylindrical housing having two ends closed by a
first and a second bracket, respectively; a first rotation member
concentrically disposed within said housing and supported by said first
bracket of the housing such that it is rotatable only in said
predetermined rotational direction; an output shaft rotatably supported by
said second bracket of the housing; a second rotation member
concentrically disposed within said housing and connected integrally to
said output shaft and rotatably supported by said second bracket of the
housing, said first and second rotation members opposing each across an
axial length within said housing; an annular torsional resilient member
connecting the first and the second cup-shaped member across said axial
length, said torsional resilient member exerting between the first and the
second rotation members a torsional torque proportional to a relative
rotational displacement of the first rotation member with respect to the
second rotation member; an input shaft driven in said predetermined
rotational direction and concentrically extending within said first and
second rotation members to be rotatably supported by said first and second
rotation members; and a torque transmission mechanism for transmitting
torque from the input shaft to the first rotation member while the
torsional torque exerted by said torsional resilient member from the first
to the second rotation member in said predetermined rotational direction
is below a predetermined magnitude, said torque transmission mechanism
limiting under the predetermined magnitude the torque transmitted from the
first to the second rotation member.
According to one aspect of this invention said torque tranmission mechanism
comprises: a first internal gear formed integrally with said first
rotation member to extend axially therefrom toward the second rotation
member; a second internal gear formed integrally with said second rotation
member to extend axially therefrom toward the first rotation member; a
first sun gear supported concentrically on said input shaft in axial
alignment with said first internal gear; a second sun gear supported
concentrically on said input shaft in axial alignment with said second
internal gear, wherein the first sun gear is supported on the input the
shaft such that it is rotatable only in a direction opposite to said
predetermined rotational direction while the second sun gear is fixed on
the input shaft; a plurality of first planetary gears meshing with said
first sun gear and first internal gear and rotatably supported at an equal
eccentricity with respect to the input shaft on a first carrier rotatably
supported on the input shaft, such that the first planetary gears are
capable of planetary motion around the first sun gear; a plurality of
second planetary gears meshing with said second sun gear and second
internal gear and rotatably supported at an equal eccentricity with
respect to the input shaft on a second carrier rotatably supported on the
input shaft, such that the second planetary gears are capable of planetary
motion around the second sun gear; first satellite shafts secured on and
extending axially from respective first planetary gears with an
eccentricity with respect to respective axes of rotation of the first
planetary gears; second satellite shafts secured on and extending from
respective second planetary gears with an eccentricity with respect to
respective axes of rotation of the second planetary gears, the
eccentricity of the second satellite shafts with respect to the respective
axes of the second planetary gears being equal to the eccentricity of the
first satellite shafts with respect to the respective axes of the first
satellite gears; and a disk-shaped guide member rotatably supported on the
input shaft between said first and second rotation members and having
radially extending guide slots slidably engaging with said first and
second satellite shafts, such that a torque in said predetermined
direction is transmitted via the guide member while the torque is below
said predetermined magnitude.
According to another aspect, the torque transmission mechanism comprises: a
first sun gear concentrically supported on said input shaft; a second sun
gear concentrically supported on said input shaft, wherein the first sun
gear is supported on the input shaft such that it is rotatable only in a
direction opposite to said predetermined rotational direction while the
second sun gear is fixed on the input shaft; a plurality of first
planetary gears meshing with said first sun gear and rotatably supported
at an equal eccentricity with respect to the input shaft on a first
carrier formed integrally with said first rotation member, such that the
first planetary gears are capable of planetary motion around the first sun
gear; a plurality of second planetary gears meshing with said second sun
gear and rotatably supported at an equal eccentricity with respect to the
input shaft on a second carrier formed integrally with said second
rotation member, such that the second planetary gears are capable of
planetary motion around the second sun gear; first satellite shafts
secured on and extending axially from respective first planetary gears
with an eccentricity planetary gears; second satellite shafts secured on
and extending from respective second planetary gears with an eccentricity
with respect to respective axes of rotation of the second planetary gears,
to eccentricity of the second satellite shafts with respect to the
respective axes of the second planetary gears being equal to the
eccentricity of the first satellite shafts with respect to the respective
axes of the second planetary gears; and a disk-shaped guide member
rotatably supported on the input shaft between said first and second
rotation members and having radially extending guide slots slidably
engaging with said first and second satellite shafts, such that a torque
in said predetermined direction is transmitted via the guide member while
the torque is below said predetermined magnitude.
The novel features which are believed to be characteristic of this
invention are set forth with particularity in the appended claims. This
invention itself, however, both as to its organization and method of
operation, together with further objects and advantages thereof, may best
be understood from the detailed description of the preferred embodiments
taken in connection with the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a sectional side view of a rotation transmission device according
to a first embodiment of this invention;
FIGS. 2 through 4 shows sections along the lines II--II, III--III, and
IV--IV, respectively, of FIG. 1;
FIGS. 5a, 5b, 6a and 6b are schematic operational views of the torque
limiting transmission mechanism of the device of FIG. 1 as viewed from the
direction of the arrow A, wherein FIGS. 5(a) and 5(b) show the mechanism
in a first and a second asynchronous operation mode, respectively, and
FIGS. 6(a) and 6(b) show the mechanism in the initial asynchronous and the
final synchronous operation mode, respectively;
FIG. 7 is a sectional side view of a rotation transmission device according
to a second embodiment of this invention taken along line VII--VII of FIG.
10.
FIGS. 8 through 10 show sections along the lines VIII--VIII, IX--IX, and
X--X, respectively, of FIG. 7; and
FIGS. 11a, 11b, 12a and 12b are schematic operational views of the torque
limiting transmission mechanism of the device of FIG. 7 as viewed from the
direction of the arrow A, wherein FIGS. 11(a) and 11(b) show the mechanism
in a first and a second asynchronous operation mode, respectively, and
FIGS. 12(a) and 12(b) show the mechanism in the initial asynchronous and
the final synchronous operation mode, respectively;
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
Referring first to FIGS. 1 through 4 of the drawings, the structure of the
rotation transmission device according to the first embodiment of this
invention is described.
The housing 10 of the transmission device comprises an outer hollow
cylindrical casing 11 and a first and a second bracket 12 and 13 fixed
thereto by means of through-bolts (not shown), etc., to close the two ends
thereof. The torque acting on the input shaft 40 is transmitted to the
output shaft 33 via a transmission mechanism with a torque limiting
function according to this invention, as described in detail hereinbelow.
In this embodiment, the input shaft 10 is driven in the clockwise
direction as viewed from right (the input side) in FIG. 1. In what
follows, the rotational directions are referred to as right or left as
viewed from the input side (in the direction of the arrow A in FIG. 1).
A first cup-shaped rotation member 20, consisting of a disk-shaped base
portion and an internal gear portion 21 integral therewith, is disposed
concentrically within the housing 10 to open toward the output side
(toward left in FIG. 1). The first rotation member is supported at a boss
portion 22 on the first bracket 12 via a bearing 14 having a one-way
clutch mechanism, such that the rotation member 20 is rotatable only in
the clockwise direction. The internal gear portion 21 of the rotation
member 20 is provided with integral teeth at the inner side surface
thereof.
A second cup-shaped rotation member 30, formed integrally with the output
shaft 33 to open toward the input side, is concentrically disposed within
the housing 10 to oppose the first cup-shaped rotation member 20 across an
axial length. The second rotation member 30 and the output shaft 33 are
rotatably supported, at the boss portion 32 of the rotation member 30, on
the second bracket 13 by means of a pair of bearings 15 and 16. The second
rotation member 30 comprises an internal gear portion 31 provided with
internal teeth formed on the inner side surface thereof.
A substantially annular torsional resilient member 80 bridges the
cylindrical internal gear portions 21 and 31 of the first and second
cup-shaped rotation members 20 and 30. Namely, the torsional resilient
member 80 comprises a pair of flanges 80a and 80b and is secured to the
internal gear portion 21 of the first rotation member 20 at the flange
80a, and to the internal gear portion 31 of the second rotation member 30
at the flange 80b thereof. Thus, the torsional resilient member 80 exerts
between the first and the second rotation members 20 and 30 a torsional
torque which is proportional to the relative rotational displacement of
the first rotation member 20 with respect to the second rotation member
30.
The input shaft 40, extending through the boss portion 22 of the first
cup-shaped rotation member 20, extends concentrically within the
cup-shaped rotation members 20 and 30 to the boss portion 32 of the second
rotation member 30. The input shaft is rotatably supported at the front
end (at the left in FIG. 1) thereof on the boss portion 32 of the second
rotation member 30 via a bearing 34 and at the root portion thereof on the
boss portion 22 of the first rotation member 20 via a pair of bearings 23a
and 23b.
The torque transmission mechanism for transmitting torque from the input
shaft 40 to the first cup-shaped rotation member 20 is organized as
follows.
A first sun gear 41 is supported on the input shaft 40 in axial alignment
with the first internal gear 21 via a bearing 42 provided with the one-way
clutch mechanism, such that the first sun gear 41 is rotatable only in the
counterclockwise direction (the direction opposite to the predetermined
rotational direction). On the other hand, a second sun gear 43 is fixed on
the input shaft 40 in axial alignment with the second internal gear 31.
A pair of first planetary gears 55, fixed on the planetary shafts 54
rotatably carried by the first carrier 50, mesh with the first sun gear 41
and the first internal gear 21 on the first cup-shaped rotation member 20,
such that the first planetary gears 55 are capable of planetary motion
around the first sun gear 41. The carrier 50 comprises a pair of side
plates 51a and 51b rotatably supported on the input shaft 40 via the
bearings 52a and 52b provided on the central shaft holes thereof. The
shafts 54 of the planetary gears 55 are rotatably supported by the side
plates 51a and 51b of the carrier 50 via the bearings 53a and 53b, at an
equal eccentricity with respect to the input shaft 40. The side plates 51a
and 51b of the carrier 50 are connected via a pair of rectangular plates
(the sections of which are shown in FIG. 3) to form a rectangular box
structure opening in the radial direction.
A pair of second planetary gears 65, fixed on the planetary shafts 64 and
carried by the second carrier 60, mesh with the second sun gear 43 and the
second internal gear 31 on the second cup-shaped rotation member 30, such
that the second planetary gears 65 are capable of planetary motion around
the second sun gear 43. The carrier 60 comprises a pair of side plates 61a
and 61b rotatably supported on the input shaft 40 via the bearing 62a and
62b provided on the central shaft holes thereof. The shafts 64 of the
planetary gears 65 are rotatably supported by the side plates 61a and 61b
via the bearings 63a and 63b, at an equal eccentrically with respect to
the input shaft 40. The side plates 61a and 61b of the carrier 60 are
connected via a pair of rectangular plates (the sections of which are
shown in FIG. 4) to form a rectangular box structure opening in the redial
direction.
To the respective output side ends of the shafts 54 of the first planetary
gears 55 are fixed first satellite shafts 56 to extend axially thereform
with a predetermined eccentricity with respect to the respective central
axes of the planetary gear shafts 54. Likewise, to the input side ends of
the shafts 64 of the second planetary gears 55 are fixed second satellite
shafts 66 with a predetermined eccentricity (equal to the above
eccentricity of the first satellite shafts 56 with respect to the central
axes of the first planetary gear shafts 54) with respect to the respective
central axes of the planetary gear shafts 66. Further, first satellite
rings 58 are rotatably and concentrically supported on satellite shafts 56
via bearings 57. Likewise, second satellite rings 68 are rotatably and
concentrically supported on satellite shafts 66 via bearings 67. A guide
disk 70, rotatably supported on the input shaft 40 via a bearing 71, has
four equally spaced radially extending satellite guide slots 72 (see FIG.
2), antipodal two of which slidably engage with the first satellite rings
58 and the other antipodal two of which slidably engage with the second
satellite rings 68.
The first and the second internal gears 21 and 31 have the same number of
teeth. Likewise with respect to the first and the second sun gears 41 and
43. Thus, the first and the second planetary gears 55 and 65 have the same
number of teeth.
Next, the method of operation of the above torque transmission mechanism is
described by reference to FIG. 5a, 5b, 6a and 6b which schematically show
the mechanism as viewed from the input side (from the direction of the
arrow A in FIG. 1). FIGS. 5(a) and 5(b) show the mechanism in two distinct
operational states (in two distinct asychronous modes as described below);
FIG. 6(a) shows the mechanism in the non-operating initial state and FIG.
6(b) shows the mechanism in the state where the torque limiting function
of the mechanism is in operation. In FIGS. 5 and 6, each gear is
represented by its pitch circle and the center. Since the first and the
second sun gears 41 and 43 completely overlap with each other, they are
shown in partial representations; likewise with the internal gears of the
first and the second cup-shaped rotation members 20 and 30. The guide disk
70 is also shown partially, wherein the two guide slots represented in the
figures engage with the firstt and the second satellite ring 58 and 68,
respectively. The bearings 14 and 42 having the one-way clutch function
are represented schematically so as to show their functions clearly. Thus,
the bearing 14 supports the first rotation member 20 on the housing 10 in
such a manner that the rotation member 20 is rotatable only in the
clockwise direction (standard direction X) relative to the stationary
housing 10; on the other hand, the bearing 42 supports the first sun gear
41 on the input shaft 40 in such a manner that the sun gear 41 is
rotatable only in the counterclockwise direction (the direction opposite
to the standard direction X) relative to the input shaft 40. Further, the
annular torsional member 80 is schematically shown in the form of a
helical tension spring to represent the function thereof schematically,
and clearly.
Before embarking on the description of the operation, the meanings of the
reference characters, etc., in FIGS. 5 and 6 are first summarized:
direction X is the standard rotational direction (clockwise as viewed from
the input side);
point O is the center of the input shaft 40;
points P1 and P2 are the centers of the first and the second planetary
shafts 54 and 64, respectively;
points S1 and S2 are the centers of the first and the second satellite
shafts 56 and 66, respectively;
line L1 is the radius which connects the center O and the point P1;
line M1 is the radius which connects the center O to the point S1 and which
is the central redial line of the guide slot 72 engaging with the first
satellite ring 58;
line N1 is the line which connects the points P1 and S1, i.e., the
revolutionary radius of the first satellite ring 58 around the center P1;
angle .theta.1 represents the angle .angle.OP1S1 formed by the lines L1 and
N1, i.e., the rotational angle of the first planetary gear 55 (or the
revolutional angle of the first satellite ring 58) with respect to the
raduis L1 connecting the sun center O to the planetray center P1, wherein
the angle .theta.1 is measured positive in the direction of rotation of
the first planetary gears 55 around the center P1 (i.e., in the
counterclockwise direction), such that .theta.1 falls between -180.degree.
and +180.degree..
line L2 is the radius which connects the center O to the point P2;
line M2 is the radius which connects the center O to the point S2 and which
is the central radial line of the guide slot 72 engaging with the second
satellite ring 68;
line N2 is the line which connects the points P2 and S2, i.e., the
revolutionary radius of the second satellite rings 68 around the center
P2;
angle .theta.2 represents the angle .angle.OP2S2 formed by the lines L2 and
N2, i.e., the rotational angle of the second planetary gear 65 (or the
revolutional angle of the second satellite ring 68) with respect to the
radius L2 connecting the sun center O to the planetary center P2, wherein
angle .theta.2 is measured positive in the direction of rotation of the
second planetary gears 65 around the center P2 (counterclockwise), such
that .theta.2 falls between -180.degree. and +180.degree..
In the following description of the method of operation of the transmission
mechanism according to the above embodiment, the mechanism is said to be
in a synchronous state when the above-defined angles .theta.1 and .theta.2
are equal in magnitude and sign (.theta.1=.theta.2); the mechanism is said
to be in an asynchronous state, when the angles .theta.1 and .theta.2 are
different from each other (.theta.1.noteq..theta.2). It is characteristic
of the transmission mechanism according to this invention that the modes
of operation are different according to whether the mechanism is in a
synchronous or an asynchronous state, and, when it is in an asynchronous
state, according to whether cos .theta.1 is greater than cos .theta.2 (cos
.theta.1>cos .theta.2), referred to as the first asynchronous mode, or cos
.theta.1 is smaller than cos .theta.2 (cos .theta.<cos .theta.2), referred
to as the second asynchronous mode.
Referring first to FIG. 5(a), the first asynchronous mode operation of the
transmission mechanism is described, where cos .theta.1 is greater than
cos .theta.2. In the state shown in FIG. 5(a), let us make the fundamental
assumption that:
(1) the output shaft 33 and hence the second rotation member 30 remains
stationary. Further let us make (at least for the present) a further
assumption (2) that
(2a) the first rotation member 20 also remains stationary, and that
(2b) the first sun gear 41 rotates at the same speed with the second sun
gear 43 (i.e., the first sun gear 41 is stationary relative to the input
shaft 40).
Under these assumption (1) and (2), the first and the second planetary
gears 55 and 65 rotates counterclockwise around the respective centers P1
and P2 thereof and revolves clockwise around the center O of the input
shaft 40, the rotational and revolutional velocities of the first and
second planetary gears 55 and 65 being equal to each other, respectively.
Thus, the revolutional velocities .omega.1 of the centers P1 and P2 of the
first and the second planetary gears 55 and 65 around the center O are
equal to each other. However, the revolutional velocities of the centers
S1 and S2 of the satellite rings 58 and 68 around the sun center O are not
equal to each other. Namely, since the planetary gears 55 and 65 rotate
around the respective centers P1 and P2 thereof and hence the centers S1
and S2 of the satellite rings themselves revolve around the respective
centers P1 and P2, the clockwise revolutional velocity of the satellite
center S1 or S2 around the sun center O is obtained by adding to the
revolutional velocity .omega.1 of the planetary center P1 or P2 around the
sun center O the revolutional velocity component .omega.2 of the satellite
center S1 or S2 with respect to the sun center O, which component results
from the revolution of the satellite center S1 or S2 around the planetary
center P1 or P2. Namely, the total angular revolutional velocity .omega.
of the satellite center S1 or S2 around the sun center O is given by:
.omega.=.omega.1+.omega.2, where the revolutional velocity component
.omega.2 is proportional to cos .theta.1 or cos .theta.2 and inversely
proportional to the length of the radius OS1 or OS2. Since the lengths of
the radii OS1 and OS2 are substantially constant and equal to each other,
the velocity components .omega.2 of points S1 and S2 can be regarded to be
substantially proportional to the magnitudes of cos .theta.1 and cos
.theta.2, respectively. Thus, the revolutional velocity components
.omega.2 of the points S1 and S2 vary periodically with the passage of
time.
If the mechanism is in the first asynchronous state where cos .theta.1 is
greater than cos .theta.2 (cos .theta.1>cos .theta.2) as shown in FIG.
5(a), and if the above assumptions (1) and (2) are both maintained, the
total revolutional angular velocity of the point S1 around the center O
becomes greater than that of the point S2 around the center O. This
consequence, however, is impossible. Namely, the points S1 and S2 lie on
the central radial lines M1 and M2, respectively, of the guide slots 72 of
the guide disk 70, and hence the angle .angle.S1OS2 should remain constant
(equal to a right angle in the case of the embodiment). Thus, if the
assumption (1) is maintained, either (2a) or (2b) of the second assumption
(2) should be discarded. In the case shown in FIG. 5(a), the revolutional
velocity of the point S1 should be further reduced to be made equal to
that of the point S2. This can be effected either by reducing the
rotational velocity of the first sun gear 41 or by rotating the first
rotation member 20 in the counterclockwise direction. The latter however
is impossible due to the function of the one-way clutch of the bearing 14.
The former, on the other hand, is possible in view of the free-rotational
direction of the one-way clutch of the bearing 42. Thus, the assumption
(2b) should be discarded.
Thus the method of operation of the mechanism in the first asynchronous
mode may be summarized as follows: When the input shaft 40 is driven, the
second planetary gears 65 are driven via the second sun gear 43 fixed on
the input shaft 40. The movements of the second planetary gears 65 entail
the movements of the second satellite rings 68. The guide disk 70 engaging
with the satellite rings 68 is thus rotated in the clockwise direction in
accordance with the revolutional and rotational velocity of the second
planetary gears 65. Due to the constraint acting on the first planetary
gears 55 from the guide disk 70 via the first satellite rings 58, the
first sun gear slips in the counterclockwise direction relative to the
input shaft 40 while the first rotation member 20 remains stationary due
to the action of the one-way clutch bearing 14. Thus, during the first
asynchronous mode, the torsion of the torsional resilient member 80
remains constant while the difference between the angles .theta.1 and
.theta.2 is reduced.
FIG. 5(b) represents the transmission mechanism in the second asynchronous
mode where cos .theta.1 is less than cos .theta.2 (cos .theta.1<cos
.theta.2). If we make the same above-mentioned assumptions (1) and (2), an
argument similar to the above leads to an impossible conclusion that the
revolutional velocity of the first satellite centers S1 around the sun
center O in the clockwise direction is smaller than that of the second
satellite centers S2. As described above, the angle formed by the lines
OS1 and OS2 is fixed--equal to a right angle--since the lines OS1 and OS2
are the central radial lines of the guide slots 72 of the guide disk 70.
In order to increase the revolutional velocity of the first satellite
centers S1 around the sun center O, either the rotational velocity of the
first sun gear 41 should be increased, or the first rotation member 20
should be rotated in the clockwise standard direction X. In view of the
one-way clutch function of the bearings 14 and 42, only the latter
alternative is possible. Namely, if the first assumption (1) is
maintained, we are forced to discard the first (2a) of the second
assumption (2).
Thus the method of operation of the mechanism in the second asynchronous
mode where cos .theta.1<cos .theta.2 may be summarized as follows: When
the input shaft 40 is driven, the second planetary gears 65 are driven via
the second sun gear 43 fixed on the input shaft 40. The movements of the
second planetary gears 65 entail the movements of the second satellite
rings 68. The guide disk 70 engaging with the satellite rings 68 is thus
rotated in the clockwise direction in accordance with the revolutional and
rotational velocity of the second planetary gears 65. Due to the
constraint acting on the first planetary gears 55 from the guide disk 70
via the first satellite rings 58, the first rotation member 20 rotates in
the clockwise direction relative to the second rotation member 30 while
the first sun gear 40 is stationary relative to the input shaft 40 and
rotates together therewith, due to the one-way clutch function of the
bearing 42. Thus, during the second asynchronous mode, the angular
displacement of the first rotation 20 member with respect to the second
rotation member 30 increases, thereby increasing the torsional torque of
the torsional resilient member 80 acting between the first and second
rotation members 20 and 30; further, the difference between the angles
.theta.1 and .theta.2 is reduced.
Thus, both in the first and the second asynchronous modes, the first
planetary gears 55 rotate in the clockwise direction relative to the
second planetary gears 65, thereby reducing the difference between the
angles .theta.1 and .theta.2. Consequently, if the mechanism is in the
asynchronous state (i.e., .theta.1.noteq..theta.2), the operation of the
mechanism continuously reduces the difference between the angles .theta.1
and .theta.2, the mechanism thereby tending toward the synchronous state
(.theta.1=.theta.2).
FIG. 6(a) shows the mechanism in the initial non-operating state. Namely,
the angular displacement of the first rotation member 20 relative to the
second rotation member 30 is null, and hence the resilient member 80
exerts null torsional torque between the first and the second rotation
members 20 and 30. On the other hand, the angle .theta.1 is greater than
the .theta.2 (.theta.1>.theta.2), and the mechanism is in an asynchronous
state. Thus, when the input shaft 40 is driven in the clockwise standard
direction X, the mechanism is operated alternately in the first and the
second asynchronous modes, as described above by reference to FIGS. 5(a)
and 5(b). Namely, if the mechanism is initially in the first asynchronous
mode, then it passes into the second asynchronous mode, and alternates
between the first and the second asynchronous mode operations thereafter.
During the first asynchronous mode periods, the rotational displacement of
the first rotation member 20 relative to the second rotation member 30,
and hence the torsional torque exerted by the torsional resilient member
80 therebetween, remains constant. On the other hand, during the second
asynchronous mode periods, the first rotation member 20 is rotated in the
clockwise direction relative to the second rotation member 30 such that
the torsional torque exerted by the torsional resilient member 80
increases. Consequently, the torsional torque of the resilient member 80
is accumulated.
The output shaft 33 integral with the second rotation member 30 is thus
acted on by an increasing output torque equal to the torsional torque
exerted by the resilient member 80 from the first rotation member 20 to
the second rotation member 30. The reaction of the output torque acts
partially on the input shaft 40. However, it is grounded in the main part
thereof on the housing 10 via the one-way clutch mechanism of the bearing
14. When the increasing output torque thus increases to become equal to
(or greater than) the torque acting on the output shaft 33 from an
exterior load (not shown), the output shaft 33 begins to rotate together
with the load. Otherwise, the output torque increases until the mechanism
finally reaches the synchronous state where the output torque takes its
predetermined maximum.
FIG. 6(b) shows the mechanism in the final synchronous state in which the
output torque is at the predetermined maximum and the torque limiting
function according to this invention is in operation. In the final
synchronous state, the rotational displacement of the first rotation
member 20 relative to the second rotational member 30 is at the maximum,
such that the resilient member 80 exerts a predetermined maximum torsional
torque from the first rotation member 20 to the second rotation member 30.
The reaction of the output torque in this state acts on and is grounded by
the housing 10 in its totality via the one-way clutch function of the
bearing 14, rather than acting on the input shaft 40. Since the mechanism
is in the synchronous state (.theta.1=.theta.2), the above assumptions (1)
and (2) ((2a) and (2b)) can be satisfied simultaneously. Thus, the
relative position of the first and the second rotation members 20 and 30
remains the same, while all the other gears of the mechanism continue to
rotate without transmitting torque therebetween. Once the mechanism
reaches the synchronous torque limiting state, it is retained in such
state unless the load torque acting on the output shaft 33 becomes less
than the predetermined maximum torque.
As described above, although the predetermined maximum torque acts on the
output shaft 33 in the final synchronous torque limiting operation mode,
the reaction on the input shaft 40 is null in principle. Thus, so long as
the output shaft 33 remains stationary, the motive power or energy is not
dissipated, in principle, regardless of how fast or how long the input
shaft is rotated.
Referring next to FIGS. 7 to 10 of the drawings, the structure of the
rotation transmission device according to a second embodiment of this
invention is described.
The housing 10 of the transmission device comprises an outer hollow
cylindrical casing 11 and a first and a second bracket 12 and 13 fixed
thereto by means of through-bolts (not shown), etc., to close the two ends
thereof. The torque acting on the input shaft 40 is transmitted to the
output shaft 33 via a transmission mechanism with a torque limiting
function according to the second embodiment of this invention, as
described in detail hereinbelow. As in the case of the first embodiment,
the input shaft 40 is driven in the clockwise direction as viewed from
right (the input side) in FIG. 7. In what follows, the rotational
directions are referred to as right or left as viewed from the input side
(from the direction of the arrow A in FIG. 7).
A first disk-shaped rotation member 21, forming part of a first radially
open box-shaped carrier 20, is disposed concentrically within the housing
10. The first rotation member 21 is supported at a boss portion 23 thereof
on the first bracket 12 via a bearing 14 having a one-way clutch
mechanism, such that the first rotation member 21 is rotatable only in the
clockwise direction.
A second disk-shaped rotation member 31, forming part of a second radially
open box-shaped carrier 30 and formed integrally with the output shaft 33,
is concentrically disposed within the housing 10 to oppose the first
disk-shaped rotation member 21 across an axial length within the housing
10. The second rotation member 31 is rotatably supported, together with
the output shaft 33 integral therewith, on the second bracket 13 by means
of a pair of bearings 15 and 16.
A substantially annular torsional resilient member 60 bridges the
circumferences of the first and second rotation members 21 and 31 across
the axial length. Namely, the torsional resilient member 60 comprises a
pair of flanges 60a and 60b and is secured to the first rotation member 21
at the flange 60a, and to the second rotation member 31 at the flange 60b
thereof. Thus, the torsional resilient member 60 exerts between the first
and the second rotation members 21 and 31 a torsional torque which is
proportional to the relative rotational (i.e., angular) displacement of
the first rotation member 21 with respect to the second rotation member
31.
The input shaft 40, extending through the boss portion 23 of the first
rotation member 21, extends concentrically within the housing 10 through
the first and the second carrier 20 and 30, and is rotatably supported at
the front side (at the left in FIG. 7) thereof on the second carrier 30
via a pair of bearings 33a and 32b, and at the root portion thereof on the
first carrier 20 via a pair of bearings 22b and 23.
The torque transmission mechanism for transmitting torque from the input
shaft 40 to the first rotation member 21 is organized as follows.
A first carrier 20 is formed of the first rotation member 21 and a
rectangular side plate 22 which is attached integrally to the first
rotation member 21 across an axial gap via a pair of axially extending
plates to form a radially open box-structure (see FIGS. 7 and 9).
Likewise, a second carrier 30 is formed of the second rotation member 31
and a rectangular side plate 32 which is attached integrally to the
rotation member 31 across an axial gap via a pair of axially extending
plates to form a radially open box-structure (see FIGS. 7 and 10). A first
sun gear 41, disposed within the first carrier 20, is supported on the
input shaft 40 via a bearing 42 provided with the one-way clutch
mechanism, such that the first sun gear 41 is rotatable only in the
counterclockwise direction (the direction opposite to the predetermined
rotational direction) relative to the input shaft 40. On the other hand, a
second sun gear 43, disposed within the second carrier 30, is fixed on the
input shaft 40. A pair of first planetary gears 25, fixed on the planetary
shafts 24 rotatably carried by the first carrier 20, mesh with the first
sun gear 41, such that the first planetary gears 25 are capable of
planetary motions around the first sun gear 41. The shafts 24 of the first
planetary gears 25 are rotatably supported on the first carrier 20 via the
bearings 21a and 22a, at an equal eccentricity with respect to the input
shaft 40. A pair of second planetary gears 35, fixed on the planetary
shafts 34 rotatably carried by the second carrier 30, mesh with the second
sun gear 43, such that the second planetary gears 35 are capable of
planetary motion around the second sun gear 43. The shafts 34 of the
second planetary gears 35 are rotatably supported on the second carrier 30
via a pair of bearings 31a and 32a, at an equal eccentricity with respect
to the input shaft 40.
To the respective output side ends of the shafts 24 of the first planetary
gears 25 are fixed first satellite shafts 26 to extend axially therefrom
with a predetermined eccentricity with respect to the respective central
axes of the planetary gear shafts 24. Likewise, to the input side ends of
the shafts 34 of the second planetary gears 35 are fixed second satellite
shafts 36 with a predetermined eccentricity (equal to the above
eccentricity of the first satellite shafts 26 with respect to the central
axes of the first planetary gear shafts 24) with respect to the respective
central axes of the second planetary gear shafts 34. Further, first
satellite rings 28 are rotatably and concentrically supported on satellite
shafts 26 via bearings 27. Likewise, second satellite rings 38 are
rotatably and concentrically supported on satellite shafts 36 via bearings
37. A guide disk 50, rotatably supported on the input shaft 40 via a
bearing 51, has four equally spaced radially extending satellite guide
slots 52 (see FIG. 8), antipodal two of which slidably engage with the
first satellite rings 28 and the other antipodal two of which slidably
engage with the second satellite rings 38. The first and the second sun
gears 41 and 43 have the same number of teeth, and likewise the first and
the second planetary gears 55 and 65 have the same number of teeth.
Next, the method of operation of the above torque transmission mechanism
according to the second embodiment is described by reference to FIGS. 11a,
11b, 12a and 12b, which schematically show the mechanism as viewed from
the input side (from the direction of the arrow A in FIG. 7). FIG. 11(a)
and 11(b) show the mechanism in the first and the second asynchronous
modes; FIG. 12(a) shows the mechanism in the non-operating initial state
and FIG. 12(b) shows the mechanism in the state where the torque limiting
function of the mechanism is in operation. In FIGS. 11 and 12, each gear
is represented by means of its pitch circle and the center. Since the
first and the second sun gears 41 and 43 completely overlap with each
other, they are shown in partial representations. The bearings 14 and 42
having the one-way clutch function are represented schematically so as to
show their functions clearly. Thus, the bearing 14 supports the first
rotation carrier 20 on the housing 10 in such a manner that the carrier 20
is rotatable only in the clockwise direction (standard direction X)
relative to the stationary housing 10; on the other hand, the bearing 42
supports the first sun gear 41 on the input shaft 40 in such a manner that
the sun gear 41 is rotatable only in the counterclockwise direction (the
direction opposite to the standard direction X) relative to the input
shaft 40. Further, the annular torsional member 60 is schematically shown
in the form of a helical tension spring to represent the function thereof
schematically and clearly.
The meanings of the reference characters, etc., in FIGS. 11a, 11b, 12a and
12b are as follows:
direction X is the standard rotational direction (clockwise as viewed from
the input side);
point O is the sun center or the center of the input shaft 40;
points P1 and P2 are the centers of the first and the second planetary
shafts 24 and 34, respectively;
points S1 and S2 are the centers of the first and the second satellite
shafts 26 and 36, respectively;
line L1 is the radius which connects the sun center O and the planetary
center P1;
line M1 is the radius which connects the sun center O and the satellite
center S1 and which is the central radial line of the guide slot 52
engaging with the first satellite ring 28;
line N1 is the line which connects the points P1 and S1, i.e., the
revolutionary radius of the first satellite ring 28 around the planetary
center P1;
angle .theta.1 represents the angle .angle.OP1S1 formed by the lines L1 and
N1, i.e., the rotational angle of the first planetary gear 25 (or the
revolutional angle of the first satellite ring 28) with respect to the
radius L1 connecting the sun center O to the planetary center P1, wherein
the angle .theta.1 is measured positive in the direction of rotation of
the first planetary gears 25 around the center P1 (i.e., in the
counterclockwise direction), such that .theta.1 falls between -180.degree.
and +180.degree.;
line L2 is the radius which connects the sun center O to the planetary
center P2;
line M2 is the radius which connects the center O to the point S2 and which
is the central radial line of the guide slot 52 engaging with the second
satellite ring 38;
line N2 is the line which connects the points P2 and S2, i.e., the
revolutionary radius of the second satellite ring 38 around the planetary
center P2;
angle .theta.2 represents the angle .angle.OP2S2 formed by the lines L2 and
N2, i.e., the rotational angle of the second planetary gear 35 (or the
revolutional angle of the second satellite ring 38) with respect to the
radius L2 connecting the sun center O to the planetary center P2, wherein
angle .theta.2 is measured positive in the direction of rotation of the
second planetary gears 35 around the center P2 (counterclockwise), such
that .theta.2 falls between -180.degree. and +180.degree..
As in the case of the first embodiment, the mechanism is said to be in a
synchronous state when the above-defined angles .theta.1 and .theta.2 are
equal in magnitude and sign (.theta.1=.theta.2); the mechanism is said to
be in an asynchronous state, when the angles .theta.1 and .theta.2 are
different from each other (.theta.1 .noteq..theta.2). The modes of
operation are different according to whether the mechanism is in a
synchronous or an asynchronous state, and, when it is in an asynchronous
state, according to whether cos .theta.1 is greater than cos .theta.2 (cos
.theta.1>cos .theta.2), referred to as the first asynchronous mode, or cos
.theta.1 is smaller than cos .theta.2 (cos .theta.<cos .theta.2), referred
to as the second asynchronous mode.
Referring first to FIG. 11(a), the first asynchronous mode operation of the
transmission mechanism according to the second embodiment is described. In
the state shown in FIG. 11(a), let us make the fundamental assumption
that:
(1) the output shaft 33 and hence the second carrier 30 (and the second
rotation member 31) remains stationary. Further let us make (at least for
the present) a further assumption (2) that
(2a) the first carrier 20 (and the first rotation member 21) also remains
stationary, and that
(2b) the first sun gear 41 rotates at the same speed with the second sun
gear 43 (i.e., the first sun gear 41 is stationary relative to the input
shaft 40).
Under these assumptions (1) and (2), the first and the second planetary
gears 25 and 35 rotate counterclockwise around the respective centers P1
and P2 thereof at the same rotational velocity, the centers P1 and P2 of
the first planetary gears 25 and the second planetary gears 35 being
stationary. Thus, the satellite centers S1 and S2 revolve around the
stationary centers P1 and P2, respectively, at the same revolutional
velocity. Due to the revolutions around the respective centers P1 and P2,
the satellite centers S1 and S2 come to possess angular velocities around
the sun center O which are proportional to cos .theta.1 and cos .theta.2,
respectively, and inversely proportional to the length of the radius OS1
or OS2, respectively. (In the case shown in FIG. 11(a), the angular
velocity of the point S1 around the sun center O is clockwise, while that
of the point S2 around the sun center O is counterclockwise.) Since the
lengths of the radii OS1 and OS2 are substantially constant and equal to
each other, the angular velocity of points S1 and S2 around the sun center
O can be regarded to be substantially proportional to the magnitudes of
cos .theta.1 and cos .theta.2, respectively. Thus, the revolutional
angular velocity of the points S1 and S2 around the sun center O vary
periodically with the passage of time.
Thus, if the mechanism is in the first asynchronous state where cos
.theta.1 is greater than cos .theta.2 (cos .theta.1>cos .theta.2) as shown
in FIG. 11(a), and if the above assumptions (1) and (2) are both
maintained, the angular velocity of the point S1 around the sun center O
becomes greater than that of the point S2 around the sun center O. This
consequence, however, is impossible. Namely, the points S1 and S2 lie on
the central radial lines M1 and M2, respectively, of the guide slots 52 of
the guide disk 50, and hence the angle .angle.S1OS2 should remain constant
(equal to a right angle in the case of the embodiment). Thus, if the
assumption (1) is maintained, either (2a) or (2b) of the second assumption
(2) should be discarded. In the case shown in FIG. 5(a), the angular
velocity of the point S1 around the sun center O should be reduced to that
of the point S2. This can be effected either by reducing the rotational
velocity of the first sun gear 41 or by rotating the first carrier 20 in
the counterclockwise direction. The latter however is impossible due to
the function of the one-way clutch of the bearing 14. The former, on the
other hand, is possible in view of the free-rotational direction of the
one-way clutch of the bearing 42. Thus, the assumption (2b) should be
discarded.
Thus the method of operation of the mechanism in the first asynchronous
mode may be summarized as follows: When the input shaft 40 is driven, the
second planetary gears 35 are driven via the second sun gear 43 fixed on
the input shaft 40. The movements of the second planetary gears 35 entail
the movements of the second satellite rings 38. The guide disk 50 engaging
with the satellite rings 38 is thus rotated in accordance with the
rotational velocity of the second planetary gears 35. Due to the
constraint acting on the first planetary gears 25 from the guide disk 50
via the first satellite rings 28, the first sun gear 41 slips in the
counterclockwise direction relative to the input shaft 40 while the first
carrier 20 remains stationary due to the action of the one-way clutch
bearing 14. Thus, during the first asynchronous mode, the torsion of the
torsional resilient member 60 remains constant while the difference
between the angles .theta.1 and .theta.2 is reduced.
FIG. 11(b) represents the transmission mechanism in the second asynchronous
mode where cos .theta.1 is less than cos .theta.2 (cos .theta.1<cos
.theta.2). If we make the same assumptions (1) and (2), an argument
similar to the above leads to an impossible conclusion that the
revolutional angular velocity of the first satellite centers S1 around the
sun center O in the clockwise direction is smaller than that of the second
satellite centers S2. (As described above, the angle formed by the lines
OS1 and OS2 is fixed--equal to a right angle--since the lines OS1 and OS2
are the central radial lines of the guide slots 52 of the guide disk 50.)
In order to increase the revolutional velocity of the first satellite
centers S1 around the sun center O, either the rotational velocity of the
first sun gear 41 should be increased, or the first carrier 20 should be
rotated in the clockwise standard direction X. In view of the one-way
clutch function of the bearings 14 and 42, only the latter alternative is
possible. Namely, we are forced to discard the first (2a) of the second
assumption (2).
Thus the method of operation of the mechanism in the second asynchronous
mode where cos .theta.1<cos .theta.2 may be summarized as follows: When
the input shaft 40 is driven, the second planetary gears 35 are driven via
the second sun gear 43 fixed on the input shaft 40. The movements of the
second planetary gears 35 entail the movements of the second satellite
rings 38. The guide disk 50 engaging with the satellite rings 38 is thus
rotated in accordance with the rotational velocity of the second planetary
gears 35. Due to the constraint acting on the first planetary gears 25
from the guide disk 50 via the first satellite rings 28, the first carrier
20 rotates in the clockwise direction relative to the second carrier 30
while the first sun gear 40 is stationary relative to the input shaft 40
and rotates together therewith, due to the one-way clutch function of the
bearing 42. Thus, during the second asynchronous mode, the angular
displacement of the first rotation member 21 with respect to the second
rotation member 31 increases, thereby increasing the torsional torque of
the torsional resilient member 60 acting between the first and second
rotation members 21 and 31; further, the difference between the angles
.theta.1 and .theta.2 is reduced.
Thus, both in the first and the second asynchronous modes, the first
planetary gears 25 rotate in the clockwise direction relative to the
second planetary gears 35, thereby reducing the difference between the
angles .theta.1 and .theta.2. Consequently, if the mechanism is in the
asynchronous state (i.e., .theta.1.noteq..theta.2), the operation of the
mechanism continuously reduces the difference between the angles .theta.1
and .theta.2, the mechanism thereby tending toward the synchronous state
(.theta.1=.theta.2).
FIG. 12(a) shows the mechanism according to the second embodiment in the
initial non-operating state. Namely, the angular displacement of the first
carrier 20 in the clockwise direction relative to the second carrier 30 is
at the minimum (i.e., the rotational displacement of the first rotation
member 21 relative to the second rotation member 31 is null), and hence
the resilient member 60 exerts null torsional torque between the first and
the second rotation members 20 and 30. On the other hand, the angle
.theta.1 is greater than the angle .theta.2 (.theta.1>.theta.2), and the
mechanism is in an asynchronous state. Thus, when the input shaft 40 is
driven in the clockwise standard direction X, the mechanism is operated
alternately in the first and the second asynchronous modes, as described
above by reference to FIGS. 11(a) and 11(b). Namely, if the mechanism is
initially in the first asynchronous mode, then it passes into the second
asynchronous mode, and alternates between the first and the second
asynchronous mode operations thereafter. During the first asynchronous
mode periods, the rotational displacement of the first rotation member 21
(the first carrier 20) relative to the second rotation member 31 (the
second carrier 30), and hence the torsional torque exerted by the
torsional resilient member 60 therebetween, remains constant. On the other
hand, during the second asynchronous mode periods, the first rotation
member 21 is rotated in the clockwise direction relative to the second
rotation member 31 such that the torsional torque exerted by the torsional
resilient member 60 increases. Consequently, the torsional torque of the
resilient member 60 is accumulated.
The output shaft 33 integral with the second rotation member 31 is thus
acted on by an increasing output torque equal to the torsional torque
exerted by the resilient member 60 from the first rotation member 21 to
the second rotation member 31. The reaction of the output torque acts
partially on the input shaft 40. However, it is grounded in the main part
thereof on the housing 10 via the one-way clutch mechanism of the bearing
14. When the increasing output torque thus increases to become equal to
(or greater than) the torque acting on the output shaft 33 from an
exterior load (not shown), the output shaft 33 begins to rotate together
with the load. Otherwise, the output torque increases until the mechanism
finally reaches the synchronous state where the output torque takes its
predetermined maximum.
FIG. 12(b) shows the mechanism in the final synchronous state in which the
output torque is at the predetermined maximum and the torque limiting
function according to this invention is in operation. In the final
synchronous state, the rotational displacement of the first carrier 20
relative to the second carrier 30 is at the maximum, such that the
resilient member 60 exerts a predetermined maximum torsional torque from
the first rotation member 21 to the second rotation member 31. The
reaction of the output torque in this state acts on and is grounded by the
housing 10 in its totality via the first carrier 20 and the one-way clutch
function of the bearing 14, rather than acting on the input shaft 40.
Since the mechanism is in the synchronous state (.theta.1=.theta.2), the
above assumptions (1) and (2) ((2a) and (2b)) can be satisfied
simultaneously. Thus, the relative position of the first and the second
carriers 20 and 30 remains the same, while all the gears of the mechanism
continue to rotate without transmitting torque therebetween. Once the
mechanism reaches the synchronous torque limiting state, it is retained in
such state unless the load torque acting on the output shaft 33 becomes
less than the predetermined maximum torque.
As described above, although the predetermined maximum torque acts on the
output shaft 33 in the final synchronous torque limiting operation mode,
the reaction on the input shaft 40 is null in principle. Thus, so long as
the output shaft 33 remains stationary, the motive power or energy is not
dissipated, in principle, regardless of how fast or how long the input
shaft is rotated.
The principle of this invention is applicable to transmission devices other
than the first and the second embodiments described above. For example, in
the case of the above embodiments, the first sun gear 41 is mounted on the
input shaft 40 via a bearing 42 having the one-way clutch function, the
second sun gear 43 being fixed on the input shaft 40. However, the first
sun gear 41 may be fixed on the input shaft 40, the second sun gear 43
being mounted on the input shaft via a bearing having a one-way clutch
function, provided that the locking and free-rotating directions of the
one-way clutch bearing is selected appropriately. Further, the number of
the first or the second planetary gears is not limited to two. The number
of the first and the second planetary gears (and that of the first and the
second satellite rings) may be one, two, three, four or more, and the
number of the first planetary gears (or that of the first satellite gears)
may be different from that of the second planetary gears (that of the
second satellite gears). Further, in the case of the above embodiments,
the satellite shafts are secured on the ends of the planetary shafts.
However, when the eccentricity of the satellite shafts with respect to the
respective axes of the planetary shafts is selected at a greater value,
the satellite shafts may be secured on the side surfaces of the planetary
gears. Furthermore, in the case of the above embodiments, the guide slots
of the guide disk engaging with the satellite rings extend radially
straight (along the radially straight directions M1 and M2). However, the
whole or a part thereof may be curved. Further, in the case of the above
embodiments, the guide slots of the guide disk slidably engage with the
satellite shafts via annular satellite rings rotatably supported on the
satellite shafts. However, the guide slots of the guide disk may directly
and slidably engage with the satellite shafts, or, even if satellite rings
are utilized, they may take forms other than that of the annular rings,
which forms may comprise planar or curved surfaces matching with the forms
of the side surfaces of the guide slots of the guide disk. Further, the
first and the second carriers for carrying the planetary gears may take
many well-known forms, provided that they are capable of supporting the
planetary gears to allow planetary motions.
The principle of this invention, by which a gear transmission mechanism is
utilized for limiting the transmitted torque and hence no energy is
dissipated, is completely different from that of the conventional torque
limiting transmission devices in which the torque is limited by means of a
frictional sliding contact. Since the amount of heat generated by friction
is negligible, the reliability and the durability of the device is greatly
enhanced. The transmission device according to this invention thus
provides an ideal flexible joint which is applicable to a wide variety of
torque transmission systems. For example, in the relatively dynamic
application field, the device according to this invention provides an
ideal transmission in the case where electric motors or internal
combustion engines, which intrinsically have constant rpm characteristics,
are utilized as the prime mover for industrial machines or automobiles,
whose rotational speeds are to be varied over a wide range. The relatively
static application fields include those of torque multiplier devices,
winding devices, or various types of screw fastener devices.
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